Controllable hydraulic geartype machine



Feb. 19, 1963 e. WIGGERMANN CONTROLLABLE HYDRAULIC GEAR-TYPE MACHINE Filed Dec. 10, 1957 4 Sheets-Sheet l WEN Feb. 19, 1963 G. WIGGERMANN 3,077,835

CONTROLLABLE HYDRAULIC GEAR-TYPE MACHINE Filed Dec. 10, 1957 4 Sheets-Sheet 2 G. WIGGERMANN Feb. 19, 1963 4 v CONTROLLABLE HYDRAULIC GEAR-TYPE MACHINE Filed Dec. 10, 1957 4 Sheets-Sheet 3 Feb. 19, 1963 e. WIGGERMANN 3,077,835

CONTROLLABLE HYDRAULIC GEAR-TYPE MACHINE Filed Dec. 10, 1957 4 Sheets-Sheet 4 3,077,835 Patented Feb. 19, 1%63 his 342177335 C(BNTRQLLABLE HYDRAULEC GEAR- TYPE MAQHINE Georg Wiggermann, Kresshronn, Germany, assignor to Walter Reiners, Monchen-Gladbach, Germany Filed Dec. 19, 1%7, Ser. No. 701,764 Claims priority, application Germany Dec. 11, 1956 6 Claims. (tCi. 103-3) My invention relates to hydraulic positive-displacement machines suitable as pumps or motors, and in a more particular aspect to rotary hydraulic machines of the gear type.

As a rule, gear-type pumps or motors, compared with positive-displacement machines of other types, excel in simplicity of design, reliability of operation, compactness, and relatively low manufacturing cost. Inherently, however, gear-type machines have constant volumetric delivery and constant delivering direction for a given speed and direction of rotation. It has, therefore, been proposed to broaden the fields of use by making geartype machines controllable at least with respect to volumetric delivery. The main features of the prior proposals are devices that permit varying the effective width of the gear teeth. In practice, however, such machines were found unsuitable or unfavorable ecause of their intricacy and high cost of manufacture.

it is an object of my invention to devise gear-type hydraulic machines that aiford varying and controlling the volumetric delivery per rotation and, if desired, also the delivering direction for a given direction of gear rotation, while preserving the above-mentioned fundamental advantages of the conventional, non-controllable gear-type machines.

To this end, and in accordance with one of the features of my invention, I provide the gears of the machine with involute teeth and make the axial spacing of the two coacting gears of a gear pair continuously variable, preferably with the aid of an eccentric on which one of the gears is journalled. These and other features of my invention will be explained presently with reference to underlying considerations that are closely related to the fundamental functioning of gear-type hydraulic machines in general. Reference will be had to the explanatory il ustration in FIG. 7 of the drawing showing schematically a conventional gear-type machine and indicating the terms used in the following definitions and equations.

As a rule, positive-displacement machines of the gear type are provided with a pair of intermeshing spur gears Z and Z which are enclosed at their axial sides as well as at the peripheries by a housing H leaving only enough clearance to permit a gliding and sealing fit. The inlet and outlet openings I and O of the housing communicate with the gear chamber of the housing H at the two sides respectively of the gear-meshing range.

f one looks upon the gear pair from one of the front sides, as shown in FIG. 7, and imagines the pair bisected by a geometric plane P-P passing vertically through the axes of the respective spur gears, then the teeth of the spur gears may be thought to pass through the plane PP in the range of three rectangular windows, the middle one of these three windows being located in the meshing range and traversed by the teeth of both gears Z and Z The volumetric delivery per rotation of the machine can be determined from the sum of the volumes-formed by the totality of tooth structure and hydraulic medium-that pass through these windows during a single full rotation of the driven gear, provided the positive and negative signs of the volumes at the respective three windows are chosen in accordance with the flow direction; the total volume of flow through the middle window common to both gears being given a negative sign, and the flow volume through each of the other two windows being thought to be positive.

Used in the following are the terms defined presently (see FIG. 7): a=spacing between the respective axes of the two gears =axial width of the gear teeth F F =tooth cross section, seen frontally, of respective gears Z Z Z1, z =number of teeth of gears Z Z dg dg =root-circle diameter of gears Z Z dw dw =pitch-circle diameter of gears Z Z dk dk =addendum-circle diameter of gears Z 2 The volume flowing through the imaginary windows of the plane PP can be calculated, with suillcient accuracy, as representing a hollow cylinder whose inner diameter corresponds to dg, whose outer diameter corresponds to dk, and whose height corresponds to the width b of the gear teeth. Relative to the middle window common to both spur gears in the meshing range, the flow volume passing through this window can be imagined to be formed of two hollow cylinders whose inner diameters correspond to dg and zig respectively, whose outer diameters correspond to dw and dw respectively, and whose height in each case is equal to b. On this basis, the entire volumetric delivery (V) per rotation is determined by the following equation.

For any given tooth design, the values dk and dk the number of teeth Z1 and Z2, as well as the axial space mg a are known so that in each particular case the cor- In general, the two spur gears Z Z have the same dimensions so that:

2 :1 dk =dk dw =dw and the Equation 1 becomes simplified to:

V=bvr/2(dk dw (6) With both spur gears having identical dimensions: a=dw. Hence the Equation 6 can be written as:

V=b.1r/2(dk -a (7) In accordance with my invention, I draw from Equation 6 the conclusion that the delivering quantity V can be varied by varying the pitch-circle diameter dw. Hence,

it is one of the features of my invention, based upon the Equation 7, to obtain in gear-type displacement machines a variation of the delivering quantity by varying the axial spacing of a pair of intermeshing spur gears.

Such a delivery control requires of the gear teeth that the desired variation in axial spacing can be carried out during operation and without disturbance of the correct transmission of driving power between the gears. While in the conventional, non-controllable gear-type displacement machines any tooth design is suitable to a greater or lesser extent as long as it receives a power transmission of substantially constant transmission ratio, it is another feature of my invention to provide the delivery-controllable machines with involute teeth because such teeth permit a limited change of center-to-center distance without impairing the desired uniform power transmission at constant velocity ratio and the mutual rolling movement of the teeth flanks.

In order to achieve a largest possible range of volumetric control, it is further desirable to utilize all of the variation in center spacing afforded by the involute teeth, and to design these teeth so as to extend this range as much as feasible. According to my invention, therefore, the addendum-circle diameter dk and the rootcircle diameter a'g as well as the number of teeth z of the involute spur gears are so chosen that (1) When the axial spacing between the two gears is a minimum, the add circles (dk) intersect the then obtaining meshing line at respective points that are located in the vicinity of, or are identical with, the points of contact between the meshing line and the respective root circles;

(2) When the axial spacing is a maximum, the degree of overlap of the respective spur gears (meshing degree) has still at least the value 1;

(3) The tips of the gear teeth are relatively pointed so that the ratio between the largest tooth thickness and the thickness at the tip (measured in arcuate degrees) is approximately 6 to 8. That is, the peripheral width at the tip is only about /6 to A: of the largest width of each tooth.

As is the case with conventional, non-controllable geartype hydraulic machines, the two spur gears in a controllable machine according to the invention must be in continuous sealing engagement with the housing over a certain extent of the gear periphery, so that the hydraulic liquid cannot flow back from the pressure side along the gear periphery to the suction side. For that reason, a machine according to the invention requires means that satisfy the just-mentioned sealing requirenints but neverthelesspermit a center-to-center displacement of at least one of the two gears relative to the other.

According to a further feature of my invention, therefore, at least one of the spur gears of a pair of intermeshing gears is journalled on an'eccentric member which, for stepless variation of the axial spacing and hence of the volumetric delivery, is rotatably mounted inthe machine housing. According to a more specific feature, the eccentric member is provided with a portion that surrounds part of the periphery of the eccentrically journalled gear with tight running fit; and this portion of the eccentric member has a cylindrical back surface concentric to the rotation axis of the eccentric member and in sealing engagement with a cylindrical wall portion of the housing that is likewise concentric to the rotation axis of the eccentric. Thus, the eccentric member forms a hydraulic seal together with the housing in all available rotational positions of the eccentric member. i i

The foregoing and other objects, advantages and features of my invention will be apparent from, and will be mentioned in, the following description of the embodiments illustrated by way of example on the accompanying drawings, in which FIG. 1 is an axial section through a two-gear displacement machine;

FIG. 2 is a cross section through the same machine;

FIG. 3 is an axial section of a three-gear displacement machine;

FIG. 4 is a front view of the machine shown in FIG. 3 with one portion of the housing removed;

FIG. 5 is a cross-sectional view of another three-gear displacement machine;

FIG. 6 illustrates schematically an assembly composed of two two-gear displacement machines and appertaining control devices;

FIG. 7 is a diagrammatical and explanatory illustration for the purpose of definition as already set forth above; and

FIGS. 8 and 9 are further explanatory diagrams relating to modifications of machines according to the invention.

The controllable two-gear positive-displacement machine illustrated in FIGS. 1 and 2 comprises a housing formed of a main portion 1 and a cover portion 2, in which the two spur gears Z and Z are journalled in meshing engagement with each other. Only a few teeth of each gear in the meshing range are illustrated in FIG. 2. The gear Z is keyed to the driving shaft 3 of the machine and is directly journalled on the housing structure. The short shaft 4 of spur gear Z is journalled in an eccentric member 5 which is journalled at its outer diameter within a cylindrical recess of housing portions 1 and 2. The eccentric member 5 has a shaft 6 passing through the cover portion 2 and protruding outwardly therefrom. The protruding end of shaft 6 carries a control arm 8 secured to shaft 6 by means of a cotter pin 7. The eccentric member 5 encloses the spur gear Z at its two axial sides and also at its periphery with a close running fit. The large diameter of the eccentric member 5 is journalled in the wall of housing 1, 2 with such a close running fit that a return flow of the oil or other hydraulic medium from the pressure side to the suction side of the machine is prevented in all angular positions of the eccentric member.

The hollow space 9 in the housing surrounds the spur gear Z over the major portion of the gear periphery with a close sealing fit. However, in the range of spur gear Z the interior space 9 of the housing is radially widened to such an extent that the spur gear Z can readily perform the displacement transverse to its axis occurring when the control shaft 6 of member 5 is turned by means of arm 8 within the entire available range of control displacement.

The two gears have involute teeth so designed as to permit a largest possible variation in center spacing a (FIG. 1) between the two gears. For this purpose, the root-circle diameter, the addendum-circle diameter and the number of teeth are so chosen relative to one another that the teeth are almost pointed at the tip, and that with minimum axial spacing or virtually clearance-free meshlng engagement, the addendum circles of the gears intersect the meshing line substantially at those points where the meshing line runs tangential to the root circles; furthermore, at the largest axial spacing, the addendum circles cut from the then obtaining meshing line a portion that is approximately equal to the meshing division, i.e. the degree of overlap e is at least equal to the value one. When a particularly great control range is desired for operation with only moderate delivering pressure, the center spacing between the spur gears may be reduced to such an extent that the teeth become undercut near the root.

The hydraulic medium to be delivered enters into the housing through the bore 10 and leaves the housing through a bore 10a. Both bores open into the interior 9 of the housing and are coaxialiy aligned at opposite sides of the meshing range.

The operation of the machine for pumping action is as follows. The spur gear Z is continuously driven to rotate clockwise (FIG. 2) and to drive the spur gear 2;, counterclockwise. The tooth gaps of both spur gears fill themselves with liquid medium entering through the bore ill. The liquid is entrained between the teeth toward the left-hand side of the housing space 9. From the pressure side, the medium leaves the machine through bore ltta. At the same time, the teeth in the meshing range of both gears prevent the medium from returning from the pressure side at the left to the suction side at the right of the housing, with the exception of a residual quantity determined by the inevitable dead space between the teeth in the meshing range of the two gears.

The control or regulation of the volumetric delivery of the gear-type displacement machine, here described for operation as a pump, is predicated upon the fact that the just-mentioned dead space, acting in the negative sense upon the volumetric delivery, is varied by changing the spacing a between the spur gears Z and Z This change in spacing is produced by turning the eccentric member, thus displacing the shaft 4 of gear Z relative to the axis of the driving shaft 3.

In the illustrated example of FIGS. 1 and 2, the eccentricity of member 5 is such that when the axis of rotation of respective gears Z and Z and of the eccentric 5 are located on a straight line viewed from the front (FIG. 2), the two gears mesh nearly without clearance. With this setting, the control arm 8 is located in the outermost left-hand position as determined by a stationary stop 8a. Since now the dead spaces in the meshing range are a minimum, this control position of arm 8 corresponds to the maximum (V of volumetric delivery per rotation.

On the other hand, when control arm 8 is shifted toward the right into the outermost position determined by another stop 8b, then the spur gears Z and 2 are set by the eccentric member into the kinematically largest possible spacing, corresponding to an overlapping or meshing range of the gear teeth approximately equal to one. Consequently, now the dead spaces in the meshing range are a maximum. The resulting maximum return flow of hydraulic medium from the pressure side to the suction side results in the minimum (V of positive delivery.

Consequently, the control and regulating range of the machine extends continuously between the volumetric quantities V and V and can be varied in a stepless manner by changing the angular setting of the control arm 8. This range of control can be made to extend approximately from one-half up to full delivery.

According to another object of my invention, however, the control range can be extended down to zero delivery; and it is also an object of my invention to permit a stepless control or regulation of delivery from the full value in one direction to the full value in the opposite direction.

For achieving these further objects, and in accordance with other features of my invention, I provide the machine with at least three spur gears of which one is simultaneously in mesh with the two other gears and forms an intermediate gear; and l journal this intermediate gear on an eccentric member rotatable in the machine housing so that in one limit position the intermediate gear Z has the largest available spacing from one of the two other gears but the smallest spacing from the third gear, whereas in the other limit position the intermediate gear has smallest spacing from the first-mentioned other gear but largest spacing from the third gear. The two gears meshing with the intermediate gear are both journallecl in the machine housing so that their geometric axes are fixed relative to the housing. Furthermore, the eccentrically displaceable intermediate gear Z is rotatable in the housing with considerable radial clearance relative thereto, whereas the two other gears Z Z have a major portion of their respective peripheries enclosed by the housing with a close running and sealing fit. As will be explained presently with reference to the embodiment shown in FIGS. 3 and 4, such a machine permits a stepless regulation of the delivering quantity within a control range extending from zero delivery up to the maximum in each of two delivering directions.

The machine shown in FIGS. 3 and 4 has a housing formed of a main portion 21 and a cover portion 22. Rotatably mounted in the housing are three spur gears Z Z Z which are continuously in meshing engagement with each other. The spur gear Z is keyed to a driving shaft 3 which is journalled in the housing so that the geometric axes of shaft 3 and spur gear Z are fixed relative to the housing structure. The spur gear Z is seated on a short shaft 23 which is likewise journalled in the housing in fixed axial relation thereto. The spur gears Z and Z are closely surrounded by the housing along the major portions of the respective gear peripheries as will as at the axial sides of the gears so as to form a sealing fit. The intermediate spur gear Z is journalled on an eccentric member 24 which has coaxially aligned shafts 25 and 25a journalled in the housing. The spur gear Z is radially spaced from the inner Walls of the housing. The shaft 25 of eccentric 24 passes through the cover portion 22 of the housing and may be provided on its protruding end with a hand wheel or other control member corresponding to the control arm 8 described above with reference to FIGS. 1 and 2.

The eccentricity c of eccentric member 24 and the locations of the axes of rotation of gears Z and Z respectively relative to the housing are so chosen that, for example in the illustrated condition, the center-tocenter spacing between gears Z and Z is a maximum while simultaneously the spacing between gears Z and Z is a minimum. Rotation of eccentric member 24 results in a stepless and mutually inverse variation of the two spacings and eventually, upon rotation of angular degrees, sets the machine for minimum center spacing between gears Z and Z As regards the type and design of the teeth, the principles explained with reference to FIGS. 1 and 2 are also applicable to the machine of FIGS. 3, 4. The hydraulic medium is supplied into the hollow space (FIG. 4) through inlet and outlet bores 10 and 16a axially aligned and located on opposite sides of the intermediate gear Z The operation is as follows. When the spur gear Z is driven clockwise, gear Z rotates in the same sense and intermediate gear Z in the opposite sense. It follows that the volume entrained by the tooth gaps of the respec tive gears Z and Z along the housing wall can be neglected in the further consideration because these two volumes cancel each other. The actual delivering quantity of the displacement machine is determined by the dead spaces obtaining in the respective two meshing engagements formed by the intermediate gear Z with the respective gears Z and Z Relative to delivering action, these two groups of dead spaces have always mutually opposed eifects; and an effective delivery comes about only when the dead spaces in the two respective meshing engagements are made to differ from each other. The available possibility of varying the two axial spacings a and a in inverse relation to each other by means of the eccentric 24 thus permits setting the machine to an operating condition in which a =a so that the dead spaces in the two meshim engagements equal each other. That is, the delivering quantity can be controlled down to the zero value.

When the eccentric 24- is turned away from the justmentioned zero position, there will occur a difference between the amounts of spacing a and a and this is accompanied by a corresponding diiference in the volume of the dead spaces within the respective two meshing engagements so that an increasing or decreasing volumetric delivery will occur. Furthermore, the resulting delivering direction, relative to a given direction of rotation of the driving gear Z is dependent upon the direction in which the eccentric 24 is turned away from the zero position. This is so because the spur gear Z becomes displaced upwardly or downwardly depending upon the sense of rotation of the eccentric (FIG. 4).

If this regulating operation is observed, for instance, with reference to the left portion of the housing space 26 communicating with the bore 10a, then the highest position of the eccentric shown in FIG. 4 results in smallest dead spaces in the upper meshing engagement of gear Z and in largest dead spaces in the lower meshing engagement. Consequently, in this case there is a delivering excess from the left-hand portion of the housing space to the right-hand side communicating with the bore 10. On the other hand, when the spur gear Z by turning the eccentric 24, is displaced downwardly, the delivery is first reduced to zero. Further rotation of the eccentric increases the dead spaces between Z and Z and simultaneously reduces the spaces between Z and Z As a result, now an increase in excess delivery to the left-hand side of the housing space 26 and hence to the bore 10a will occur.

Consequently, a geantype displacement machine with three spur gears and two meshing engagements, in which the amounts of center spacing of the intermediate gear from the respective two other gears is variable in two inversely related directions, can be controlled in a stepless manner from zero-delivery up to maximumdelivery in either direction of delivery.

If the design of a three-gear machine according to the invention is such that, in axial view as shown in FIG. 4, the center points of gears Z and Z and the axis of rotation of the eccentric are located on a straight line, then the control device can be given the smallest possible eccentricity e of the eccentric member. In this case, the entire positive and negative control range corresponds to a rotation of the eccentric through 180 angular degrees.

A rotation through such a wide angular range is often difiicult to obtain by means of control levers and connecting rods. According to another feature of my invention, however, this difficulty is avoided by mounting the three gears in the housing of the machine in such a manner that the respective gear center points form a triangle, and by giving the eccentric member in the housing such a position and eccentricity that in the middle or zero position of the eccentric the same amounts of center spacings obtain relative to the two respective meshing engagements, whereas in the two limit positions of the eccentric the smallest possible spacing occurs at one meshing engagement and the largest possible spacing at the other.

With such a machine design, the required angle of rotation of the eccentric is equal to the angle defined by the respective center points of gears Z and Z together with the rotation axis of the eccentric, the latter forming the apex of the angle. In this manner, it is readily possible to limit the total displacement range of the eccentric member to an angle of 90.

These features are embodied in the machine illustrated in FIG. 5. The machine is a three-gear positive-displacement device whose performance generally corresponds to that described above with reference to FIGS. 3, 4. Hence, it will be suificient to mainly refer to the essential differences. The axes of gears Z Z and the rotation axis of the eccentric 24 form a triangle. This results not only in reduced over-all height of the machine as compared with a machine according to FIGS. 3, 4, but also reduces the angle of rotation required of the eccentric member for controlling or regulating the delivery from zero to maximum in both delivering directions.

The torque transmitted by the intermediate gear Z for example from Z to Z and the liquid pressure imposed upon the intermediate gear in the transverse direction, impart torques to the shaft of the eccentric member. For facilitating the control or regulating performance, these torques are preferably kept as small as feasible. For this reason, a smallest feasible eccentricity e is preferred. Such minimum eccentricity is obtained if, at minimum center spacing, the center points of Z or Z as well as the center point of Z and the corresponding rotation center point of the eccentric, areall located on a straight line as is the case in the embodiment shown in FIGS. 3, 4. However, as explained, this has'the consequence that a relatively large angle of rotation of the eccentric is required so that difiiculties of linking the control shaft of the eccentric to other devices may be encountered. Now, in the embodiment according to FIG. 5, the rotation axis of the eccentric forms a right angle with the respective axes of the spur gears Z and Z Consequently, with a smallest possible eccentricity e=a -a the entire displacing range of the control arm 8 extends through an angle of only In cases where the force or moment required for displacing the control shaft of the eccentric is irrelevant, a smaller displacing angle of the control shaft down to approximately 90 can also be achieved with an embodiment generally according to FIGS. 3 and 4, simply by increasing the eccentricity e of the eccentric member 24.

In some cases it is desired to make the machine selfregulating as regards the dependence of its volumetric delivery upon the delivering pressure. This requirement can readily be satisfied by providing pressure-responsive means which automatically turn the eccentric member and the spur gear Z in dependence upon, or opposition to, an elastic counterforce which varies with the rotation angle of the eccentric. In the latter case, the abovementioned intentional enlargement of the eccentricity e has the advantage of improving the sensitivity of the automatic regulation by increasing the torque available for the automatic control displacement.

Relative to such automatic regulation in dependence upon hydraulic pressure, particularly favorable conditions are afforded by machines of the type shown in FIGS. 1 and 2. Such two-gear machines make it possible, by suitable choice of the eccentricity e and by virtue of the large diameter of the eccentric member 5, to obtain virtually any desired regulating characteristic of continuous performance.

As mentioned, a controllable gear-type machine with a single pair of spur gears cannot be controlled down to zero-delivery but has a control range from maximum delivery down to approximately one half of that delivery. This limitation, however, can be eliminated by mechanically interconnecting the power shafts of two two-gear machines, hydraulically connecting both machines to a pair of common hydraulic conduits, and interposing a hydraulic switch-over valve between at least one. of the two machines and the conduits. By properly correlating the delivery control of the two machines and of the switch-over valve or valves, preferably by means of a common controlling or programming device, such a machine set can be controlled or regulated through any desired range of delivery from a maximum down to zero or, if desired, from maximum in one direction through zero to maximum in the other direction.

A hydraulic machine set of the type just described is schematically illustrated in FIG. 6. Two individually controllable two-gear machines A and B, each designed in accordance with FIGS. 1 and 2, have their respective input shafts aligned and connected with each other so that both are driven simultaneously. The hydraulic lines 27 and 27 extending from the respective bores 10a and 10 (FIG. 2) of machine A are connected through a hydraulic switching device, formed of a five-way slide valve 28 (FIG. 6), with the common hydraulic lines 29 and 30 of the machine set. Analogously, the hydraulic lines 127 and 127 coming from respective bores 10a and 10 of machine unit B are connected through another hydraulic switching valve 28' with the same lines 29 and 30. As explained, such machinery affords performing the same controlling operations as the above-described three-gear machines. This is so because the hydraulic switching devices 28 and 28 permit connecting the pressure and suction sides of each unit either with line 29 and line 30, or reversely with line 30 and line 29. Hence, in each delivering direction, approximately the range between one-half and full delivery can be adjusted in each unit by correspondingly displacing the respective control arms 3 and 8, while both machine units are operating in the same delivering direction relative to the common hydraulic lines 29 and 30. The control range extending from one-half delivery in each delivering direction down to zero delivery is likewise obtained by setting the control arms 3 and 8, but in this case the switching valves 28 and 28 are so set relative to each other that the two machines A and B operate opposed to each other so that their respective deliveries will cancel each other partially or, in the case of zero-delivery, entirely.

In cases where delivery in a single direction is suflicient, one of the two hydraulic switching devices 28, 28' may be omitted. Delivery control from zero to maximum is then obtained as follows. First the eccentrics and the switching valve are set by a suitable control device so that the deliveries of the two units are equal but op posed so as to cancel each other. Hence, the assembly as a whole has zero-delivery. Then, for obtaining increasing delivery, the two units are controlled, by varying the setting of the respective eccentrics, so as to decrease the delivery of the negatively operating unit down to the minimum while simultaneously increasing the delivery of the positively operating unit up to the maximum. Thereafter, the setting of the hydraulic switching device is changed to switch the previously negative unit to positive delivery while simultaneously setting the eccentric of the always positively operating unit down to approximately smallest delivery. From that point on, both eccentrics are adjusted in any desired sequence or simultaneously up to the maximum delivery of both units.

As mentioned, the controllable three-gear machines permit obtaining a similar control performance, Without the aid of accessory equipment, simply by turning a single eccentric member for passing through the entire range of control. Although a machine set composed of at least two two-gear pairs according to FIG. 6 involves greater requirements as to material and space, the possibility of operating both machine units A and B in parallel permits obtaining a somewhat greater maximum delivery in both directions of rotation for a given available over-all space and also permits the use of a manufacturers stock of twogear machines for a variety of purposes including control from maximum to one half, maximum to zero, and maximum to reversed maximum.

In the above-described embodiments of machines according to the invention, the gears are designed as spur gears of external type. However, a control according to the invention, requiring a variation in spacing between inter-meshing spur gears, can also be carried out with the aid of an internal or annular gear cooperating with an external gear, it being of no consequence, in principle, whether the external gear or the internal gear is radially displaceable within the housing of the machine.

In the example schematically shown in FIG. 8, two external gears Z Z mesh with an internal gear Z which surrounds the two external gears and corresponds, as regards functioning, to the intermediate gear Z in a three-gear machine ot the type described above with reference to FIGS. 3 and 4. In order to obtain the same control performance as in the above-described machines, the annular gear 2 may be fixed relative to the machine housing and the two external spur gears Z Z may be displaceable transverse to the axis of the annular gear while maintaining a fixed mutual center spacing a or the two external gears Z and Z may be journalled in the housing so as to have fixed geometric axes relative thereto, and the internal gear Z may be displaceable transversely thereto. In either case, a displacement between the annular gear on the one hand and the pair of external gears on the other hand causes the center spacing between the annular gear and one of the external gears to become larger while, inversely, the center spacing between the annular gear and the other external gear becomes smaller.

The spur gears in hydraulic machines according to the invention may be provided with inclined or helical teeth. Herringbone gears are preferably suitable. It is particularly advantageous to compose each individual gear of two parts coaxially located one beside the other and, if desired, firmly joined with each other, both having the same axial Width and being provided with helical teeth of mutually opposed pitch directions so that the composite gear forms herringbone teeth. Such gears reduce noise and balance the axial thrust resulting "from the two groups of inclined teeth.

Furthermore, the invention is not limited to machines with planar types of spur gears. The gears Z Z or Z Z Z may also be of the spacial type. In this case, mhey are preferably provided with conical teeth whose apexes coincide in a single point. The eccentricity e in such gears is given by the fact that the rotation axis of the center-displaceab le gear Z or of the eccentric forms an acute angle with the rotation axis of the eccentriccontrol shaft. Furthermore, in this case, the spur gears have spherical shape at their axial sides and their periphery is conical. To secure the necessary sealing, the abutting or sealing faces between the housing and the gears are given a corresponding mating shape.

Hydraulic machines according to the invention may readily be provided with the customary load-relieving grooves in the range adjacent to the efiective meshing engagement, such grooves having the purpose to avoid unnecessary compression of the oil between the meshing teeth.

For simplifying the mechanical drive, it is preferable to always apply or take ofif driving power from the shaft of a gear that has a fixed journalling axis relative to the machine housing, as is the case in all embodiments of the invention herein illustrated and described.

It will be apparent to those skilled in the art, upon studying this disclosure, that my invention permits of various modifications and may be embodied in machine designs other than those specifically illustrated and described herein, without departing from the essential features of my invention and within the scope of the claims annexed hereto.

I claim:

1. A hydraulic positive-displacement machine, comprising a housing having inlet and outlet ducts, three gears rotatable in said housing and having involute teeth, one of said gears being in meshing engagement with said two other gears, said gears forming together with said housing respective chambers communicating with said ducts, an eccentric member rotatably mounted in said housing, said one gear being journalled on said member in eccentric relation to the rotation axis of said member, said two other gears having respective axes of rotation fixed relative to said housing whereby rotary displacement of said member causes said one gear to inversely vary the spacing of its axis from the axes of said other gears respectively in a direction transverse to a line between said latter axes, said hou-sing having radial clearance relative to said one gear and having a sealing fit relative to the axial sides and periphery of said two other gears.

2. In a controllable hydraulic machine according to claim 1, said gears having respective addendum circles whichat minimum spacing between the respective axes of said gears-intersect the line of meshing engagement at respective points substantially coincident with the points where said meshing line touches the respective root circles, and said gears havingat maxim-um spacing a meshing degree at least of unity value, the peripheral thickness of the gear teeth at the tip of each tooth being approximately one sixth to one eighth of the largest tooth thickness measured in angular degrees.

3. In a controllable hydraulic machine according to claim 1, said three gears having their respective axes of rotation located substantially on a straight line, and said member having an eccentricity approximately equal to,

11 but slightly smaller than, the dedendum of said teeth, whereby rotation of said member through 180 causes a change in delivery between positive and negative maximum values.

4. In a controllable hydraulic machine according to claim 1, said three gears having their respective axes of rotation located substantially on a triangle, said eccentric member having a mid-position in which the axis of rotation of said one gear has equal amounts of spacing from the two axes of said respective other gears to obtain zero delivery of the machine, and said member having two limit positions in which said two amounts of spacing are-in inverse relation to each othera maximum and a minimum respectively, the total amount of rotary displacement of said member being at most about 90.

5. A hydraulic positive-displacement machine, comprising two gear-type units, each of said units having a housing with inlet and outlet ducts, two involute-teeth gears rotatable in said housing and meshing with each other, said gears and said housing jointly defining two chambers communicating with said respective ducts, one of said gears in each unit being displaceable toward and away from the rotation axis of the other gears through a displacement range within which said two gears remain meshed, and each unit having a control member opera-tively connected with said one gear for displacing it through said range, the other gears of said respective units being mechanically interconnected to rotate to gether with each other, a hydraulic inlet line and an outlet line, conduit means connecting said lines with said ducts of said respective units, and switching valve means interposed between at least one of said units for selectively connecting said ducts of said unit to said lines, whereby both units are selectively operable with mutually opposed and mutually uniform directions of delivery respectively. i

6. In a controllable hydraulic machine according to claim 1, said gear-s having helical teeth.

References'Cited in the file of this patent UNITED STATES PATENTS 1,602,740 Bechler Oct. 12, 1926 1,691,713 Frey Nov. 13, 1928 1,704,704 Grant Mar. 12, 1929 2,457,465 Grosser Dec. 28, 1948 2,526,964 Muller Oct. 24, 1950 2,549,241 Rorive Apr. 17, 1951 2,754,765 Ioy July 17, 1956 2,948,228 Ahlen -2 Aug, 9, 1960 FOREIGN PATENTS 250,580 Great Britain Aug. 12, 1926 297,770 Great- Britain a July 18, 1929 599,654 Great Britain Mar. 17, 1948 280,136 Italy Nov. 29, 1930 472,159 Switzerland z May 1, 1935 1,195,288 France May 19, 1959 

1. A HYDRAULIC POSITIVE-DISPLACEMENT MACHINE, COMPRISING A HOUSING HAVING INLET AND OUTLET DUCTS, THREE GEARS ROTATABLE IN SAID HOUSING AND HAVING INVOLUTE TEETH, ONE OF SAID GEARS BEING IN MESHING ENGAGEMENT WITH SAID TWO OTHER GEARS, SAID GEARS FORMING TOGETHER WITH SAID HOUSING RESPECTIVE CHAMBERS COMMUNICATING WITH SAID DUCTS, AN ECCENTRIC MEMBER ROTATABLY MOUNTED IN SAID HOUSING, SAID ONE GEAR BEING JOURNALLED ON SAID MEMBER IN ECCENTRIC RELATION TO THE ROTATION AXIS OF SAID MEMBER, SAID TWO OTHER GEARS HAVING RESPECTIVE AXES OF ROTATION FIXED RELATIVE TO SAID HOUSING WHEREBY ROTARY DISPLACEMENT OF SAID MEMBER CAUSES SAID ONE GEAR TO INVERSELY VARY THE SPACING OF ITS AXIS FROM THE AXES OF SAID OTHER GEARS RESPECTIVELY IN A DIRECTION TRANSVERSE TO A LINE BETWEEN SAID LATTER AXES, SAID HOUSING HAVING RADIAL CLEARANCE RELATIVE TO SAID ONE GEAR AND HAVING A SEALING FIT RELATIVE TO THE AXIAL SIDES AND PERIPHERY OF SAID TWO OTHER GEARS. 